The present invention relates to a low axial stiffness thrust bearing and more particularly to a thrust bearing which may be used in combination with a conventional thrust bearing on a single or coupled shaft system.
Prior art thrust bearings such as a tilting pad or tapered land bearing, have a high axial stiffness under load, generally measured in millions of pounds per inch. A typical use of such system is to axially locate the drive shaft of a turbine-generator set such as that illustrated schematically in FIG. 1. As shown by way of example in FIG. 1, the drive shaft 10 of a gas turbine 12 is coupled to the central shaft 14 of an electric generator 16 by a clutch mechanism 18. The drive shaft 10 is radially supported by a pair of radial bearings 20 and the central shaft 14 is supported by a pair of radial bearings 22. During normal operation, clutch 18 couples shafts 10 and 14 such that gas turbine 12 drives electric generator 16. In this mode, electric generator 16 generates electric power which may be applied to a utility power grid or other electrical system. When clutch 18 is engaged, shafts 10 and 14 operate as a single shaft which is axially aligned by a standard hydrodynamic thrust bearing 24 such as the tilting pad or tapered land types. As long as clutch 18 remains engaged, the single thrust bearing 24 serves to axially align both drive shaft 10 and central shaft 14 with respect to gas turbine 12 and electric generator 16, respectively.
In those applications wherein electric generator 16 is supplying a utility power grid, it is often desireable to have clutch 18 disengage in order that generator 16 may run as a motor for electrical network power factor correction. In such cases, and in the absence of other provisions, the axial alignment of central shaft 14 is no longer maintained by thrust bearing 24 and shaft 14 is free to move axially due to the magnetic forces which result from electric generator 16 being electrically "off-center." The simplest way to solve this problem would appear to be to place an additional thrust bearing between electric generator 16 and clutch 18. While such a thrust bearing would insure that central shaft 14 remains axially aligned with respect to electric generator 16, the use of an additional standard thrust bearing is impractical for the following reasons.
The interiors of the turbine and generator contain high heat sources such as hot gases, steam or current carrying electrical conductors. The exterior turbine-generator shells, base and foundation, however, are in a relatively cool ambient environment. Two points on the shaft separated by a given axial distance will therefore separate more from thermal expansion than two equally distant points on the base. Thus if a standard design thrust bearing were placed between generator 16 and clutch 18, as illustrated schematically in FIG. 3 by thrust bearing 46, the distance between the thrust collars 26 and 48 would increase during a turbine-generator start-up. The axial distance between the two base-mounted bearing support points of thrust bearings 24 and 46, however, would grow a lesser amount. This difference in axial thermal expansion together with the extremely high stiffness of a standard hydrodynamic thrust bearing would create destructive thrust loads at the two thrust bearings. It is clear therefore that two standard thrust bearings cannot be used on a single or coupled shaft system in such a situation.
In order to avoid this problem, the prior art has designed complicated clutch mechanisms which will enable thrust bearing 24 to absorb the thrust on central shaft 14 imparted by electric generator 16 during the intervals in which clutch 18 disengages shafts 10 and 14. A simplified schematic diagram of one such prior art clutch is illustrated in FIG. 2. As shown therein, the primary components of the clutch 18 is an axially displaceable housing 28, a pair of thrust collars 30, 32 and sliding sleeve 34. The housing 28 is mounted in a stationary housing 36 in a manner which permits housing 28 to move in an axial direction but prevents the rotation of housing 28 about the axis of shafts 10 and 14. Thrust collars 30, 32 are coupled to shafts 10, 14, respectively, for rotation therewith. A projection 38 on each thrust collar 30, 32 is received in a corresponding recess 40 formed in housing 28 so as to define respective thrust bearings.
The sliding sleeve 34 is located radially inward of thrust collars 30, 32 and is slidable in the axial direction. A plurality of teeth 42 are formed about the outer perimeter of opposite ends of sliding sleeve 34 and engage corresponding teeth 44 located on the inner periphery of the inner ends of thrust collars 30, 32. When clutch mechanism 18 is engaged, sliding sleeve 34 is in the position illustrated causing thrust collars 30, 32 to rotate as a single unit. Additionally, any thrust forces placed on central shaft 14 by electric generator 16 are transmitted to drive shaft 10 via axially slidable housing 28 due to the inter connection between thrust collars 30, 32 and housing 28. As a result, any thrust forces placed on shaft 14 are absorbed by thrust bearing 24.
When clutch mechanism 18 is disengaged, sliding sleeve 34 is moved axially to the right as viewed in FIG. 2 so as to disengage teeth 42, 44. In this condition, thrust collars 30, 32 (and with them shafts 10, 14) are free to rotate independently of one another. However, thrust forces placed on central shaft 14 are still transmitted to shaft 10 via axially slidable housing 28 due to the inter connection between thrust collars 30, 32 and housing 28.
When the clutch mechanism 18 is disengaged and drive shaft 10 is stationary, any thrust load placed on central shaft 14 will ultimately be absorbed by the thrust bearing 24. Relative rotary sliding motion, however, will exist between the projection 38 on the thrust collar 30 and the corresponding recess 40 in the housing 28. These relatively sliding surfaces of projection 38 and recess 40 comprise a thrust bearing which must be capable of transmitting the full thrust load from rotating central shaft 14 to stationary drive shaft 10.
The foregoing arrangement provides for the absorption of thrust load from central shaft 14 when the clutch 18 is either engaged or disengaged. It also avoids destructive thrust loads arising from thermal expansion differences. However, the clutch 18 must be designed to include a thrust bearing consisting of projection 38 and recess 40. Both the clutch 18 and the thrust bearing designs are complicated and compromised by their interdependence. Thus such a clutch mechanism is likely to require more frequent repair than if the thrust bearing function is separated from the clutch function.